This invention relates to engines such as those used to power garden tractors, lawnmowers, sump pumps, portable generators, snow blowers and the like. More particularly, this invention relates to a counterbalance weight system for canceling the primary and higher order vibrational forces in a single cylinder engine resulting from piston reciprocation.
A major cause of vibration in a single cylinder engine is piston reciprocation. The piston is started and stopped twice during each rotation of the crankshaft, and reactions to the forces which accelerate and decelerate the piston are imposed upon the engine body as vibration in directions generally parallel to the piston axis. In installations such as garden tractors, lawnmowers and the like, the engine produces a vibration that is transmitted through the machine to the operator. This vibration is uncomfortable and could produce operator fatigue. Even in an installation where there is no element of operator fatigue such as sump pumps or portable generators, engine vibration is undesirable because it causes maintenance problems and tends to reduce the useful life of the machine.
To some extent such vibrations can be decreased by providing the engine with a counterweight fixed on its crankshaft, and located at the side of the crankshaft axis directly opposite the crankpin by which the piston, through the connecting rod, is connected to the crankshaft. More commonly, two counterweights may be used on the crankshaft, one located on each side of the piston axis. In either case, such a crankshaft counterweight arrangement produces a net resultant centrifugal force vector that is diametrically opposite to the crankpin.
If the mass of the crankshaft counterweights is great enough, their net force vector parallel to the piston axis cancels the acceleration and deceleration primary forces on the piston assembly. Such a force canceling condition exists when the crankshaft counterweights are of such mass and radius of gyration that their combined centrifugal force totally cancels the centrifugal force due to the rotating masses of the crankpin and the crank end of the connecting rod. When the counterweights are even larger, the vector component parallel to the cylinder axis of the counterweights' additional combined centrifugal force completely offsets the primary (first order) acceleration and deceleration forces of the piston. Such counterweighting may be called a condition of 100% overbalance.
Unfortunately, centrifugal force due to the crankshaft counterweights also has a component transverse to the piston axis. This produces lateral vibration. If the mass of the crankshaft counterweights produces 100% overbalance, the vibration transverse to the piston axis is excessive for practical purposes. For this reason most single cylinder engines incorporate crankshaft counterweights having a mass that provides a condition of about 50% overbalance, so that the centrifugal force due to the counterweight overbalance has a component along the cylinder axis that is equal to about 50% of the acceleration and deceleration forces on the piston assembly. This represents a compromise between the severe vibration in directions parallel to the piston axis that would result with the condition of no overbalance, and the severe vibration transverse to the piston axis that would result with the condition of 100% overbalance. With this compromise condition of about 50% overbalance, there is of course some vibration parallel to the cylinder axis and some vibration transverse to it. Unsatisfactory as it is, the use of crankshaft counterweights to provide condition of about 50% overbalance is a typical balance system utilized with small commercially available engines of this type.
Other prior art systems have one or more counter-rotating balance shafts and associated counterbalance weights in addition to the crankshaft counterweights. With this type of system, the combination of the crankshaft counterbalance weights and the counter-rotating balance shaft or shafts provides forces that cancel the primary piston inertia forces, without creating undesirable forces transverse to the cylinder axis.
Unfortunately, the counter-rotating balance shaft methods do not balance the higher order forces, and in fact some designs introduce torsional or rocking forces on the engine. When the net force of all counterbalances is not located on the piston axis, force couples remain and vibrate the engine rotationally.
Rocking forces are not generated when two counterbalancing shafts are used as depicted in FIGS. 1A and 1B, with one shaft being located on each side of the piston axis. The crankshaft counterbalance weights are just sufficient to balance the crankpin and the large end of the connecting rod, but provide no balancing for the piston and the upper end of the connecting rod. Primary piston forces are balanced by the counter-rotating weights, but this balancing is not perfect. The main reason that the piston forces are not totally balanced is that the piston forces are not truly sinusoidal, while the counter-rotating balancer forces are sinusoidal. Because of the nature of the slider-crank mechanics, the piston forces are highest near Top Dead Center (TDC), cross zero when the piston is located about halfway in the cylinder bore, and have an intermediate (reversed) value when the piston is near Bottom Dead Center (BDC).
Prior art balancing methods typically do not take into account the non-sinusoidal nature of the piston forces. Although prior art methods may counterbalance the primary forces, they typically do not entirely counterbalance higher order forces, as depicted in FIGS. 2A and 2B. FIG. 2A is a graph depicting the piston force curve and the balancer force curve in a typical prior art one-cylinder engine with counter-rotating balance shafts. FIG. 2B depicts the net force curve of the piston and balancer forces in FIG. 2A.
The net forces are essentially sinusoidal in nature and have a predominant frequency twice that of the primary forces. These "secondary" forces have lower values than the primary forces, but they are high enough to cause vibration-related failures and operator discomfort. There are also forces of a higher order than the secondary forces, but these higher order forces are of such a low magnitude that they may be disregarded.
Other prior art methods use oscillating counterbalance schemes. Since it is an oscillating mass (the piston) that produces the vibration, single cylinder engines have for a number of years been balanced by a mass oscillated in a direction opposite to the piston movement. U.S. Pat. No. 3,457,804 to Harkness discloses such a system. This system has the advantage of minimizing unwanted lateral vibrations, and does not require additional gearing in the engine. While it is theoretically possible to balance higher order vibrations by this method, it requires that the ratio of connecting rod length to eccentric radius of the balancer be the same as that of the engine. Typically, the oscillating counterbalance has been made quite heavy and given a very short stroke to keep the engine compact. This configuration makes it impractical to give the balancer the same ratio of connecting rod length to crank (eccentric) throw as the connecting rod to crank throw ratio. The result is that the oscillating counterbalance has little cancellation of higher order forces since its oscillating motion is nearly sinusoidal.